Hydromechanical transmission and control method

ABSTRACT

Methods and systems for controlling a hydromechanical transmission are proposed. In one example, a control method for a hydrostatic unit of a hydromechanical variable transmission (HVT) is presented, comprising controlling the hydrostatic unit via a feedforward control architecture including a non-linear, multi-coefficient model, wherein the hydrostatic unit comprises a hydrostatic pump and a hydrostatic motor. A desired differential pressure of the hydrostatic unit or a desired hydraulic pump displacement may be used as inputs for the model, where the model&#39;s output is a pressure difference for a pump control piston coupled to a swash plate of the hydrostatic unit. Use of the non-linear model permits the hydrostatic unit to be controlled based on load, speed, and/or torque, thereby increasing the adaptability of the control system.

TECHNICAL FIELD

The present disclosure relates generally to a hydromechanicaltransmission, and more specifically, to a feedforward control strategyfor the hydromechanical transmission.

BACKGROUND AND SUMMARY

Hydromechanical transmissions enable performance characteristics (e.g.,efficiency, shift quality, drive characteristics, control response, andthe like) from mechanical and hydrostatic transmissions to be blended tomeet certain design objectives. Certain hydromechanical transmissions,referred to in the art as hydromechanical variable transmissions (HVTs),provide continuously variable gear ratios. Hydromechanical transmissionsmay be particularly desirable due to their efficiency. Vehicles used inindustries such as agriculture, construction, mining, material handling,oil and gas, and the like have therefore made use of HVTs.

Prior HVTs have included a hydrostatic unit with a hydrostatic pump andmotor. In some designs the hydrostatic pump is a variable displacementpump. In this design, a displacement of the pump depends on an angle ofa swash plate of the hydrostatic unit, which may be adjusted by acontroller. Specifically, the controller may adjust a position of adouble-action pump control piston to control the angle of the swashplate, via a pressure of the hydraulic fluid.

In certain prior HVT control strategies, the controller directlycontrols the position of a pump control piston via a hydraulic valvewithin a mechanical feedback control loop. In this system, thedifferential pressure of the pump does not affect the control pistonbehavior, because the load effect is automatically compensated by thehydraulic valve with the mechanical feedback. In this solution thecontrol of pump displacement is simple and straightforward, but thedifferential pressure is not directly controllable.

In other prior HVT control strategies, the controller directly controlsa pressure of the hydraulic fluid on the control piston chambers,providing a complex and non-linear characteristic between the pressureon the control piston, the hydrostatic differential pressure, pumpdisplacement, and pump speed. This solution is more flexible, but theidentification and usage of the non-linear characteristic is complex andhas been oversimplified, with a consequent inaccuracy.

Further in prior control strategies, the hydrostatic pump and motor thatmay be controlled in a speed control mode, where the pump displacementis controlled based on a desired speed of the hydrostatic motor, or atorque control mode, where the pump displacement is controlled based ona desired torque of the hydrostatic motor. However, previous controlstrategies for transmission hydrostatic control units have been unableto address both the torque and speed control modes and efficientlyswitch between them, which may reduce transmission efficiency and/orconstrain transmission performance (e.g., a control lag may causeinstability and/or oscillation, over-control may cause control ringingor under-damped oscillation, etc.). For example, in U.S. Pat. No.8,762,014 B2, Lister et al teaches a strategy for controlling thehydrostatic unit solely in a torque control mode. Further, in theapproach taken by Lister, the hydrostatic transmission control strategyoversimplifies transmission kinematics, which may result in controlinaccuracies.

In one embodiment, at least a portion of the abovementioned issues maybe addressed by a control method for a hydrostatic unit of ahydromechanical variable transmission (HVT), comprising controlling thehydrostatic unit via a feedforward control architecture including anon-linear, multi-coefficient model, wherein the hydrostatic unitcomprises a hydrostatic pump and a hydrostatic motor. In some cases, thecoefficients of the multi-coefficient model may be calibrated using anautomatic calibration procedure at a late stage in manufacturing of thetransmission and/or at predetermined operating intervals. A desireddifferential pressure of the hydrostatic unit or a desired hydraulicpump displacement may be used as inputs for the model, where the model'soutput is a pressure difference for a pump control piston coupled to aswash plate of the hydrostatic unit. Using a non-linear model forfeedforward hydrostatic unit control in this manner enables thehydrostatic unit to be both efficiently and accurately controlled. Useof the non-linear model further permits the hydrostatic unit to becontrolled based on load, speed, and/or torque, thereby increasing theadaptability of the control system.

In another example, the method may further include switching between atorque control mode and a speed control mode of a hydrostatic unit basedon vehicle operating conditions; wherein the multi-coefficient model isused to determine feedforward outputs in both the speed and torquecontrol modes. In the torque control mode, a desired differentialpressure may be used as an input for the multi-coefficient model whilein the speed control mode, a desired hydraulic pump displacement may beused as an input for the model. Thus, the benefits of both usingmechanical feedback control loop and controlling a pressure of thehydraulic fluid may be achieved, where the hydrostatic differentialpressure may be controlled by compensating the pump displacement, or thepump displacement may be controlled by compensating the hydrostaticdifferential pressure. In this way, a common model may be used in boththe speed and torque control modes by strategically varying the model'sinputs in the different modes. Consequently, the control system canrapidly switch between the modes depending on vehicle operatingconditions which may reduce control latency and more generally enhancetransmission performance. Further, using a common model in both thespeed and torque control modes enables processing resources in thesystem to be conserved, thereby increasing controls system efficiency.

In yet another example, the control system may use junction filters forthe inputs of the non-linear model to ensure continuity when switchingbetween the torque and speed control modes.

It should be understood that the summary above is provided to introducein simplified form a selection of concepts that are further described inthe detailed description. It is not meant to identify key or essentialfeatures of the claimed subject matter, the scope of which is defineduniquely by the claims that follow the detailed description.Furthermore, the claimed subject matter is not limited toimplementations that solve any disadvantages noted above or in any partof this disclosure.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of a hydromechanical variabletransmission and control system.

FIG. 2 is a schematic representation of a hydrostatic unit of ahydromechanical variable transmission.

FIG. 3A is a force diagram illustrating forces on a pump control pistonin a first configuration.

FIG. 3B is a force diagram illustrating forces on a pump control pistonin a second configuration.

FIG. 3C is a force diagram illustrating forces on a pump control pistonin a third configuration.

FIG. 4A is a diagram of a first feedforward control architecture.

FIG. 4B is a diagram of a second feedforward control architecture.

FIG. 5 is a flowchart that illustrates a procedure for calibratingcoefficients of a non-linear, multi-coefficient model.

FIG. 6 is a 3D surface plot showing a change in a pressure differentialbetween two sides of a pump control piston as a function of pumpdisplacement and differential pressure of a hydraulic fluid of ahydrostatic unit.

DETAILED DESCRIPTION

The methods described herein refer to a control strategy for controllingoperation of a hydromechanical variable transmission (HVT) via afeedforward control architecture including a non-linear,multi-coefficient model. An HVT of a vehicle may include a hydrostaticunit with a variable displacement hydrostatic (e.g., hydraulic) pump andmotor, used in combination with mechanical gearing in a transmissionpowered by an engine. Specifically, the hydrostatic unit may becontrolled in two different modes, between which the hydrostatic unitmay switch during operation of the vehicle. In a speed control mode, thepump displacement variable is controlled based on a desired speed of thehydrostatic motor (e.g., where a torque of the hydrostatic motor is notcontrolled). For example, when one or more gears of the HVT aredisengaged, the hydrostatic unit may be speed controlled at a firstengagement of the HVT to synchronize a desired clutch (e.g., a firstforward clutch or a backward clutch, depending on aforward-neutral-reverse (FNR) lever), or during a “blocking function”,where the one or more gears of the of the HVT are engaged to maintainthe vehicle at a standstill and memorize a load of the vehicle before asecond blocking clutch engagement, or during a “freewheel” state, tofollow a free output speed and be ready to re-engage an appropriateclutch quickly. In a torque control mode, a pump displacement variableis controlled based on a desired torque of the hydrostatic motor (e.g.,where a speed of the hydrostatic motor is not controlled). For example,the hydrostatic unit may be torque controlled when the one or more gearsof the HVT are engaged to deliver an output torque (e.g., vehicletractive effort) to the one or more wheels of the vehicle based onoperator input (e.g., pedals, FNR, etc.).

The controller of the vehicle may switch between the torque and speedcontrol modes by adjusting a position of a pump control piston coupledto a swash plate of the hydrostatic unit via a cam joint. In oneexample, a load applied to the wheels of the vehicle is converted into aload on the hydrostatic motor, generating a pressure that may bedelivered to the pump control piston via a hydraulic fluid. The pressureon the pump control piston may be directed to a first end of the pumpcontrol piston via a first pump piston valve, or to a second, oppositeend of the pump piston via a second pump piston valve. The first pumppiston valve and the second pump piston valve may be actuatedconcurrently. The first pump piston valve is actuated to apply a firstpressure to the pump control piston in a first direction, and the secondpump piston valve is actuated to apply a second pressure to the pumpcontrol piston a second direction, within a cylinder of the pump controlpiston, where the position of the pump control piston is determined by apressure differential between the first pressure and the secondpressure. One or more control strategies may be used to actuate thefirst pump piston valve and the second pump piston valve to adjust thepump control piston to a desired position. When the pump control pistonis at the desired position, an angle of the swash plate is at a desiredangle. Since the angle of the swash plate controls a displacement of thevariable displacement pump, the displacement may be controlled bycontrolling an actuation of the first pump piston valve and the secondpump piston valve.

To increase a performance of the HVT, reduce wear on parts of the HVT,and reduce a discontinuity when switching between the torque and speedcontrol modes, a control strategy comprising a feedforward controlarchitecture is proposed, based on a polynomial non-linear model for ahydrostatic unit of the HVT. FIG. 1 shows an example HVT control systemconfigured with a hydrostatic unit, such as the example hydrostatic unitof FIG. 2 . The hydrostatic unit may include a hydrostatic pump andmotor, where a pump displacement is controlled by a pump control piston.A motion of the pump control piston may be determined by forces and/orpressures on the pump control piston and a corresponding swash plate,such as the forces illustrated in the force diagrams of FIGS. 3A, 3B,and 3C. During the torque control mode, the HVT may be controlled usinga feedforward control architecture based on the polynomial non-linearmodel in a first configuration, such as the configuration shown by FIG.4A, and during the speed control mode, the HVT may be controlled using afeedforward control loop based on the same polynomial non-linear modelin a second configuration, such as the configuration shown by FIG. 4B,where the first configuration of the polynomial non-linear model and thesecond configuration of the polynomial non-linear model share the sameparameters and coefficients. The coefficients of the polynomialnon-linear model used in both configurations may be estimated via anautomatic calibration procedure, such as the calibration proceduredescribed in method 500 of FIG. 5 . The automatic calibration proceduremay solve for the coefficients sequentially by eliminating terms of thepolynomial non-linear model, as shown graphically in FIG. 6 .

Referring now to FIG. 1 , a schematic depiction of an HVT control system100 of a vehicle is shown, including an HVT 101 mechanically coupled toa power source 102 and one or more wheels 110 of the vehicle. The powersource 104 may include an internal combustion engine, electric motor(e.g., electric motor-generator), combinations thereof, and the like. Itshould be appreciated that while FIG. 1 refers to an embodiment within avehicle, in other embodiments the HVT control system 100 may not beincluded in a vehicle, and may be included in a different machine thatgenerates torque for a purpose other than propulsion.

The HVT 101 comprises a hydrostatic unit 104, which may be rotationallycoupled to a planetary gear set 120. For example, a ring gear 126 of theplanetary gear set 120 may be rotationally coupled to the power source102 via a mechanical path 106 (e.g., a crankshaft), a sun gear 122 ofthe planetary gear set may be rotationally coupled to the hydrostaticunit 104, and a carrier 124 of the planetary gear set 120 may berotationally coupled to the one or more wheels 110 of the vehicle.During operation of the vehicle, the ring gear 126 is rotated by theengine 102 at an engine speed, and a rotation of the sun gear 122 isadjusted by the hydrostatic unit 104. By adjusting the rotation of thesun gear 122 with respect to the ring gear 126, a variable amount oftorque may be generated via the carrier 124, which is transmitted to oneor more wheels 110 of the vehicle. Thus, the engine speed may be held ata constant, efficient engine speed while the torque generated by the HVT101 is increased or decreased by controlling the hydrostatic unit 104. Asystem output shaft 136 may be coupled to a driveline with a shaft,joints, etc. that may be used to carry out the power transfer betweenthe HVT 101 and one or more axles on which the one or more wheels 110are coupled.

The HVT control system 100 may include various clutches. A power sourcedisconnect clutch 130 may be arranged between the power source 102 andthe hydrostatic unit 104, to couple and decouple the power source 102from the hydrostatic unit 104. Similarly, a disconnect clutch 134 may bearranged between the hydrostatic unit 104 and the planetary gear set120, to couple and decouple the hydrostatic unit 104 from the planetarygear set 120. The disconnect clutches 130 and 134 may be dog clutches,or friction clutches, or another kind of clutch.

The planetary gear set 120 further includes a reverse drive clutch 160,a first forward drive clutch 162, and a second forward drive clutch 164.More generally, the first forward drive clutch 162 may be referred to asa first clutch or a first forward clutch, the reverse drive clutch 160may be referred to as a second clutch or a reverse clutch, and thesecond forward drive clutch 164 may be referred to as a third clutch ora second forward clutch.

Further in one example, the first forward drive clutch 162 and thereverse drive clutch 160 may be arranged adjacent to one another andhave axes of rotation that are coaxial. Specifically, in one example,the first forward drive clutch 162 and the reverse drive clutch 160 mayeach receive rotational input from the carrier 124 of the planetary gearset 120. When engaged, the first forward drive clutch 162 causes thetransmission's output to rotate in a forward drive direction.Conversely, when engaged, the reverse drive clutch 160 causes thetransmission's output to rotate in a reverse drive direction that isopposite the forward drive direction. As such, during certainconditions, the first forward drive clutch 162 and the reverse driveclutch 160 may be simultaneously engaged to place the transmission'soutput in a blocked condition in which it is held substantiallystationary.

The disconnect clutches 130 and 134 and/or the drive clutches 160, 162,and 164 may be friction clutches that each include two sets of clutchplates. The clutch plates may rotate about a common axis and may bedesigned to engage and disengage one another to facilitate selectivepower transfer to downstream components. In this way, the clutches maybe closed and opened to place them in engaged and disengaged states. Inthe disengaged state, power does not pass through the clutch. Converselyin the engaged state, power travels through the clutch duringtransmission operation. Further, the clutches may be hydraulically,electromagnetically, and/or pneumatically actuated. For instance, theclutches may be adjusted via a hydraulic piston. The adjustability maybe continuous, in one example, where the clutch may be transitionthrough partially engaged states to a fully engaged state, where arelatively small amount of power loss occurs in the clutch. However, inother examples, the clutches may be discretely adjusted.

The hydrostatic unit 104 may be controlled by a controller 112 of thevehicle. The controller 112 may include a processor 140 and memory 142.The memory 142 may hold instructions stored therein that when executedby the processor cause the controller 112 to perform the variousmethods, control strategies, diagnostic techniques, etc., describedherein. The processor 140 may include a microprocessor unit and/or othertypes of circuits. The memory 142 may include known data storage mediumssuch as random access memory, read only memory, keep alive memory,combinations thereof, etc. The memory 142 may include non-transitorymemory.

The controller 112 may receive vehicle data and various signals fromsensors positioned in different locations in the HVT 101 and/or thevehicle. The sensors may include an oil temperature sensor 170, anengine velocity sensor 172, one or more wheel velocity sensors 174,and/or other sensors of the hydrostatic unit (e.g., torque sensors,pressure sensors, swash plate angle sensor, etc.). The controller 112may send control signals to one or more actuators of the hydrostaticunit 104, for example, to adjust an output and/or direction of a flow ofhydraulic fluid through a hydrostatic pump, as described in furtherdetail below in reference to FIG. 2 and FIGS. 3A, 3B, and 3C.Additionally, the clutches 130 and 134 may receive commands (e.g.,opening or closing commands) from the controller, and actuators in theclutches 130 and 134 or actuation systems coupled to the clutches 130and 134 may adjust the state of the clutches 130 and 134 in response toreceiving the commands. For instance, the clutches 130 and 134 may beactuated via hydraulically controlled pistons, or a different suitableclutch actuator.

The HVT control system 100 may include one or more input devices 114.For example, the input devices 114 may include a pedal of the vehicle(e.g., an accelerator pedal), a control stick (e.g., aforward-neutral-reverse (FNR) lever), one or more buttons, or similartypes of control, or combinations thereof. In one example, a FNR leveris used to operate the vehicle in a forward direction or a reversedirection, and an accelerator pedal is used to increase or decrease aspeed of the vehicle. The input devices 114, responsive to driver input,may generate a transmission speed adjustment request or torqueadjustment request and a desired drive direction (a forward or reversedrive direction).

The HVT control system 100 may automatically switch between drive modeswhen demanded. For example, the operator may request a forward orreverse drive mode speed change, and the HVT 101 may increase speed andautomatically transition between one or more drive ranges associatedwith the different drive modes, as needed. The operate may requestreverse drive operation while the vehicle is operating in a forwarddrive mode. In such an example, the HVT 101 may automatically initiate ashift (e.g., synchronous shift) between the forward and reverse drivemodes. In this way, the operator may more efficiently control thevehicle, in comparison to transmissions designed for manual drive modeadjustment. It will further be appreciated that the power source 102 maybe controlled in tandem with the HVT 101. For instance, when a speedadjustment requested is received by the controller, an output speed ofthe power source 102 may be correspondingly increased.

The HVT 101 may function as an infinitely variable transmission (IVT)where the transmission's gear ratio is controlled continuously from anegative maximum speed to a positive maximum speed with an infinitenumber of ratio points. In this way, the transmission can achieve acomparatively high level of adaptability and efficiency when compared totransmission which operate in discrete ratios.

Referring now to FIG. 2 , a detailed schematic drawing of a hydrostaticunit 200 of a vehicle is shown, which may be a non-limiting example ofthe hydrostatic unit 104 described above in reference to FIG. 1 . Thehydrostatic unit functions as a variator that provides a variable outputtorque based on an applied control differential pressure, via ahydrostatic pump 202 and a hydrostatic motor 204. The hydrostatic pump202 includes a number of pistons 214 that slide within an equal numberof respective chambers 216 of a pump carrier 208, where the pump carrier208 is rotated via a pump input shaft 236 coupled to a power source 250(e.g., an engine of the vehicle, an electric motor, etc.). The equalnumber of respective chambers ride on a variable angle swash plate 220via sliding contacts, such that the range of movement of the pistons 214is set by an angle of the swash plate 220.

The hydrostatic motor 204 comprises a similar arrangement, including anumber of pistons 210 in an equal number of respective chambers 212 of amotor carrier 206, where the motor carrier 206 rotates a motor outputshaft 239. The pistons 210 of the hydrostatic motor 204 are slidablyengaged upon a fixed swash plate 238. The chambers of the pistons 214 ofthe hydrostatic pump 202 are in fluid communication with the chambers ofthe pistons 210 of the hydrostatic motor 204 via a hydraulic fluid thatfills the chambers and intervening conduits. In one example, thehydrostatic pump 202 and the hydrostatic motor 204 are coupled by afirst hydraulic conduit 270 and a second hydraulic conduit 272 throughwhich the hydraulic fluid circulates between the hydrostatic pump 202and the hydrostatic motor 204. During a first, pump mode of operation,the hydrostatic pump 202 may flow a hydraulic fluid to the hydrostaticmotor 204 via the first hydraulic conduit 270, and receive the hydraulicfluid back from the hydrostatic motor 204 via the second hydraulicconduit 272. Alternatively, during a second, motor mode of operation,the hydrostatic pump 202 may flow a hydraulic fluid to the hydrostaticmotor 204 via the second hydraulic conduit 272, and receive thehydraulic fluid back from the hydrostatic motor 204 via the firsthydraulic conduit 270. Operation of the hydrostatic unit 200 in the pumpmode and the motor mode may be controlled by an angle of the variableangle swash plate 220, where, for example, if the angle of the variableangle swash plate 220 is less than 0 (e.g., with respect to the pumpcarrier 208) the hydrostatic unit 200 may be operated in the pump mode,and if the angle of the variable angle swash plate 220 is greater than0, the hydrostatic unit 200 may be operated in the motor mode.

The variable angle swash plate 220 may be coupled to a pump controlpiston 224, which may be actuated to adjust the angle of the variableangle swash plate 220. In one example, the variable angle swash plate220 is coupled to the pump control piston 224 via a cam joint 222, wherethe angle of the swash plate 220 is adjusted as the pump control piston224 slides within a pump control piston chamber 228 in a first directionindicated by upward arrow 230 or a second direction indicated bydownward arrow 232. For example, the swash plate 220 may be rotated in afirst rotational direction indicated by arrow 234 (e.g.,counterclockwise) by sliding the pump control piston 224 within the pumpcontrol piston chamber 228 in the first direction indicated by upwardarrow 230, and the swash plate 220 may be rotated in a second rotationaldirection indicated by arrow 235 (e.g., clockwise) by sliding the pumpcontrol piston 224 within the pump control piston chamber 228 in thesecond direction indicated by downward arrow 232.

In one example, the pump control piston 224 is actuated by hydraulicpressure, where the hydraulic fluid enters the pump control pistonchamber 228 with a first pressure in the first direction indicated byupward arrow 230 via a first pump control piston valve 229, and thehydraulic fluid enters the pump control piston chamber 228 with a secondpressure in the second direction indicated by downward arrow 232 via asecond pump control piston valve 231. The first pump control pistonvalve 229 and the second pump control piston valve 231 may beproportional pressure relief valves that are continuously adjustable,where an output pressure is limited based on electrical signals receivedfrom a controller (e.g., the controller 112 of FIG. 1 ). In one example,the first pump control piston valve 229 and the second pump controlpiston valve 231 may be solenoid valves, although other suitable valvesmay be used such as pressure reducing valve, pressure relief valve,poppet valve and the like.

The hydrostatic unit may include one or more sensors which may be usedby the controller to adjust the first pressure and/or the secondpressure to control the pump control piston. For example, the pumpcontrol piston chamber 228 may include a first pressure sensor 260arranged proximate the first pump control piston valve 229, and a secondpressure sensor 262 arranged proximate the second pump control pistonvalve 231. As another example, the hydrostatic unit 200 may also includea swash plate angle sensor 264, which may estimate an angle of the swashplate, an input shaft torque sensor 266, an output shaft torque sensor268, and/or other sensors. It should be appreciated that the sensorsincluded herein are for illustrative purposes, and in other embodiments,the hydrostatic unit may include a greater or fewer number of sensors ordifferent sensors without departing from the scope of this disclosure.

As described in greater detail below in reference to FIG. 3A, theposition of the pump control piston 224 may depend on a hydrostaticdifferential pressure of the hydrostatic unit 200, an estimateddisplacement of the hydrostatic pump 202 (e.g., estimated based on themotor and pump measured speeds and estimated volumetric efficienciesunder actual conditions), a speed of the pump input shaft 236, and apressure differential between the first pressure and the secondpressure. As the variables change, forces generated on the pump controlpiston 224 may adjust the position of the pump control piston 224 toseek an equilibrium. Thus, the position of the pump control piston 224may be controlled by the pressure differential between the firstpressure and the second pressure. For example, if the first pressure isgreater than the second pressure, the pump control piston 224 may slidein the first direction indicated by upward arrow 230, causing the swashplate 220 to rotate in the first rotational direction 234.Alternatively, if the second pressure is greater than the firstpressure, the pump control piston 224 may slide in the second directionindicated by downward arrow 232, causing the swash plate 220 to rotatein the second rotational direction 235.

As the angle of the swash plate 220 is adjusted, a displacement of thepump 202 (e.g., an amount of fluid displaced by the pistons 214) varies,generating a greater or lesser volume of hydraulic fluid received ortaken from the chambers of the pistons 210. If a greater volume ofhydraulic fluid is received from the chambers of the pistons 210, anoutput speed of the output shaft 239 is increased, while if a lesservolume of hydraulic fluid received from the chambers of the pistons 210,the output speed of the output shaft 239 is decreased. Thus, the outputspeed of the hydrostatic motor 204 varies with and is controlled by theangle of swash plate 220. In this way, a controller can control theoutput speed of the hydrostatic unit by actuating the first pump controlpiston valve 229 and the second pump control piston valve 231 to adjustthe pressure differential between the first pressure on the pump controlpiston 224 and the second pressure on the pump control piston 224.

Further, a direction of rotation of the output shaft may be switched byadjusting the angle of the swash plate 220. For example, in the exampledepicted in FIG. 2 , the output shaft 239 rotates in a rotationaldirection indicated by arrow 240, the same rotational direction as theinput shaft 236 indicated by arrow 242. As the second pressure (e.g., inthe second direction indicated by downward arrow 232) increases relativeto the first pressure (e.g., in the first direction indicated by upwardarrow 230), the pump control piston 224 slides in the second direction(e.g., down in FIG. 2 ). When the pump control piston 224 reaches themidpoint of the pump control piston chamber 228 indicated by dashed line244, the swash plate 220 is adjusted to an angle of 90° relative to thepump carrier 208. At the angle of 90°, a displacement of the hydrostaticpump 202 does not vary across the number of respective chambers 216 ofthe pump carrier 208, whereby no rotation is generated in the outputshaft 239. As the pump control piston 224 passes the midpoint of thepump control piston chamber 228 indicated by dashed line 244, the angleof the swash plate 220 is adjusted further in the second rotationaldirection indicated by arrow 235, whereby a rotation is generated in theoutput shaft 239 in a direction opposite to the arrow 240. As the angleof the swash plate 220 continues to be adjusted in the second rotationaldirection indicated by arrow 235, the displacement of the pump 202(e.g., an amount of fluid displaced by the pistons 214) increases,whereby the speed of the rotation of the output shaft 239 increases,generating torque in one or more wheels of the vehicle (e.g., in theopposite direction as arrow 240).

Thus, by adjusting the first pressure and the second pressure via thefirst pump control piston valve 229 and the second pump control pistonvalve 231, the position of the pump control piston 224 may becontrolled, thereby controlling the speed of the output shaft 239. Inone example, a dynamic control strategy is employed to actuate the firstpump control piston valve 229 and the second pump control piston valve231 to adjust the speed of the output shaft 239 in response to inputfrom an operator of the vehicle in real time. In one example, thecontrol strategy dynamically generates one or more output signals toactuate the first pump control piston valve 229 and/or the second pumpcontrol piston valve 231 in accordance with a non-linear,multi-coefficient model of a feed-forward control architecture. Themulti-coefficient model may rely on one or more formulas includingvarious coefficients and variables, which may correspond to one or moreforces and pressures applied to the pump control piston 224 of thehydrostatic unit 200.

Referring now to FIG. 3A, a force diagram 300 illustrates examples ofone or more forces and pressures applied to a pump control piston 302 inaccordance with a position of a pump control piston 302 within a pumpcontrol piston chamber 306 of a hydrostatic unit of a vehicle. The pumpcontrol piston 302 and the pump control piston chamber 306 may benon-limiting examples of the pump control piston 224 and the pumpcontrol piston chamber 228 of the hydrostatic unit 200 of FIG. 2 .

The pump control piston 302 may slide within the pump control pistonchamber 306 in a first direction 316, or a second direction 318, wherethe second direction 318 is opposite the first direction 316. Theposition of the pump control piston 302 within the pump control pistonchamber 306 may determine an angle of a swash plate 310 (e.g., the swashplate 220 of FIG. 2 ), which may be coupled to the pump control piston302 via a cam joint 308, such that as the pump control piston 302 slidesin the first direction 316 or the second direction 318, the angle of theswash plate 310 is adjusted accordingly.

A number of forces may be applied to the pump control piston 302 toslide the pump control piston 302 in the first direction 316 or thesecond direction 318. The position of the pump control piston 302 withinthe pump control piston chamber 306 may be controlled by actuating afirst pump control piston valve 338 and/or a second pump control pistonvalve 339 (e.g., the first pump control piston valve 229 and/or thesecond pump control piston valve 231 of the hydrostatic unit 200 of FIG.2 ). As the first pump control piston valve 338 is actuated open (e.g.,more open), a flow of hydraulic fluid through the first pump controlpiston valve 338 in a direction indicated by arrow 316 is increased,applying a pressure P_(x1) on the pump control piston 302 in thedirection 316. Similarly, as the second pump control piston valve 339 isactuated open (e.g., more open), a flow of hydraulic fluid through thesecond pump control piston valve 339 in the direction 318 is increased,applying a pressure P_(x2) on the pump control piston 302 in thedirection 318.

The pump control piston 302 may be coupled to an end 305 of the pumpcontrol piston chamber 306 via a spring 304, with a neutral position ofthe spring 304 indicated by the dashed midpoint line 344 at a midpointof the pump control piston chamber 306. When the spring 304 is at theneutral position and the pump control piston 302 is at the midpoint ofthe pump control piston chamber 306, the swash plate 310 may beperpendicular (e.g., at a 90° angle) to the pump control piston chamber306, whereby a displacement of the hydrostatic pump is 0 (e.g., where nohydraulic flow is generated on a hydrostatic motor of the hydrostaticunit). Due to the neutral position of the spring 304 being at themidpoint of the pump control piston chamber 306, as the pump controlpiston 302 slides from the neutral position in the first direction 316,the spring 304 may apply a mechanical pressure P_(spring) on the pumpcontrol piston 302 in the second direction 318, and as the pump controlpiston 302 slides from the neutral position in the second direction 318,the spring 304 may apply a mechanical pressure P_(spring) on the pumpcontrol piston 302 in the first direction 316. Thus, when the pressureP_(x1) (e.g., exerted by oil in the pump control chamber 306) is greaterthan the pressure P_(x2), the mechanical pressure P_(spring) may beadded to the pressure P_(x2) (e.g., counteracting the greater pressureP_(x1)) due to an expansion of the spring 304. Similarly, when thepressure P_(x2) is greater than the pressure P_(x1), mechanical pressureP_(spring) may be added to the pressure P_(x1) (e.g., counteracting thegreater pressure P_(x2)) due to a contraction of the spring 304.

Additional pressures may be generated on the pump control piston 302 asa result of forces applied by a plurality of pistons of a pump carrier(not shown in FIGS. 3A-3C) on the swash plate 310. As described above inreference to FIG. 2 , a number of pistons (not shown in FIG. 3A) of thehydrostatic pump (e.g., the pistons 214 of the hydrostatic pump 202 ofFIG. 2 ) may be slidably coupled to the swash plate 310, where thenumber of pistons may apply a force F_(j_piston) the swash plate 310 ina direction indicated by an arrow 320. The force F_(j_piston) on theswash plate 310, which may depend on a piston mass, a pump shaft speed,and a variable pump displacement, may generate a pressure P_(j_piston)on the pump control piston 302. Additionally, a force F_(pDiff) may beapplied to the swash plate 310 in a direction indicated by an arrow 322,where the force F_(pDiff) is a force acting on the swash plate 310 dueto a hydrostatic differential pressure p_(Diff) of a hydraulic fluidcirculating between a hydrostatic pump and a hydrostatic motor of thehydrostatic unit. A gain from the differential pressure P_(Diff) toF_(pDiff) may depend on characteristics of the hydrostatic pump, such asa pump timing angle. In one example, the number of pistons are housedwithin a corresponding number of chambers of a pump carrier (not shownin FIG. 3A) arranged on a first side 330 of the swash plate 310, betweenthe pump control piston chamber 306 and the swash plate 310. In anotherexample, the pump carrier is arranged on a second, opposite side 332 ofthe swash plate 310, where the swash plate 310 is between the pumpcontrol piston chamber 306 and the pump carrier, and where the forceF_(j) is applied to the swash plate 310 in a direction indicated by anarrow 334 (e.g., opposite to the arrow 320), and the force F_(pDiff) isapplied to the swash plate 310 by in a direction indicated by an arrow336 (e.g., opposite to the arrow 322).

The configuration of the hydrostatic unit may change, which may resultin a change of the hydrostatic differential pressure p_(Diff), whichapplies a proportional pressure P_(pDiff) to the pump control piston302. The hydrostatic differential pressure p_(Diff) and the pressureP_(pDiff) may depend on a load applied to wheels of the vehicle, ascommanded by an operator of the vehicle. The first pump control pistonvalve 338 and the second pump control piston valve 339, respectively,may have to be actuated to adjust the pressures P_(x1) and P_(x2) basedon the pressure P_(pDiff), (e.g., on vehicle tractive effort and/ortransmission load). Therefore, the differential pressure p_(pDiff) maybe an input into the non-linear, multi-coefficient model used to controlthe pump control piston 302 along with the pressures P_(x1) and P_(x2).

The pressure P_(pDiff) may be assigned a direction based on a sign ofp_(Diff), and the pressure P_(J_piston) may be assigned a directionbased on the angle of the swash plate 310 (e.g., whether it is positiveor negative). For example, the pressure P_(J_piston) may be definedrelative to the first side 305 of the pump control piston chamber 306,where the pressure P_(J_piston) is applied in a direction indicated byarrow 312, or the pressure P_(J_piston) P may be defined relative to thesecond, opposite side 307 of the pump control piston chamber 306, wherethe pressure P_(J_piston) is applied in a direction indicated by arrow314. Thus, the position of the pump control 302 within the pump controlpiston chamber 306 may be based on a plurality of pressures, including adifference between the pressure P_(x1) and the pressure P_(x2); thepressure P_(J_piston) on the pump control piston 302; the pressureP_(spring), which may be applied against the greater of the pressureP_(x1) and the pressure P_(x2); and the pressure P_(pDiff), whichdepends on the load applied to wheels of the vehicle. FIG. 3A may depicta first equilibrium state between the plurality of pressures, where inthe first equilibrium state the pump control piston 302 is at a midpointof the pump control piston chamber 306.

Referring to FIG. 3B, a force diagram 340 illustrates examples of theone or more forces and pressures applied to the pump control piston 302in a second equilibrium state. In one example, as a result of anincrease in the pressure P_(x2) and a corresponding decrease in thepressure P_(x1), the pump control piston 302 slides within the pumpcontrol piston chamber 306 in the second direction indicated by thearrow 314, and the spring 304 is contracted beyond the neutral pointindicated by dashed line 313. As the pump control piston 302 slideswithin the pump control piston chamber 306, an angle 342 of the swashplate 310 with respect to the pump control piston chamber 306 increases,where the angle 342 is depicted in FIG. 3B between the swash plate 310and a dashed line 344 parallel to the pump control piston chamber 306.As the angle 342 of the swash plate 310 increases, a displacement of thenumber of pistons of the hydrostatic pump slidably coupled to the swashplate 310 increases, powering the hydrostatic motor of the hydrostaticunit (not shown in FIGS. 3A-3C) to rotate an output shaft of thehydrostatic unit in a first rotational direction. In FIG. 3B, thepressure P_(J_piston) defined relative to the second, opposite side 307of the pump control piston chamber 306, and is applied to the pumpcontrol piston 302 in the direction indicated by arrow 314. A negativepressure P_(pDiff) is applied to the pump control piston 302 in thedirection indicated by arrow 312, based on a sign of the differentialpressure P_(Diff). As the angle 342 of the swash plate 310 increases,the force F_(pDiff) is applied to the swash plate 310 by thedifferential pressure p_(Diff) in a direction indicated by an arrow 322,where the force F_(pDiff) is a function of the differential pressurep_(Diff) and the angle 342, and the force F_(J_piston) is applied to theswash plate 310 in a direction indicated by an arrow 320, also as afunction of the angle 342.

Referring now to FIG. 3C, a force diagram 360 illustrates examples ofthe one or more forces and pressures applied to the pump control piston302 under a third equilibrium state. As a result of an increase in thepressure P_(x1) and a corresponding decrease in the pressure P_(x2.),the pump control piston 302 slides within the pump control pistonchamber 306 in the first direction indicated by the arrow 314, and thespring 304 is expanded beyond the neutral point indicated by dashed line313. As the pump control piston 302 slides within the pump controlpiston chamber 306, the angle 342 of the swash plate 310 with respect tothe pump control piston chamber 306 increases in a negative direction,where the angle 342 is depicted in FIG. 3C between the swash plate 310and a dashed line 344 parallel to the pump control piston chamber 306.As the angle 342 of the swash plate 310 increases negatively, adisplacement of the number of pistons of the hydrostatic pump slidablycoupled to the swash plate 310 increases, powering the hydrostatic motorof the hydrostatic unit (not shown in FIGS. 3A-3C) to rotate an outputshaft of the hydrostatic unit in a second, opposite rotational directionas described above in reference to FIG. 3B. In FIG. 3C, the pressureP_(J_piston) is defined relative to the first side 305 of the pumpcontrol piston chamber 306, where the pressure P_(J_piston) is appliedto the pump control piston 302 in the direction indicated by arrow 312,based on the angle 342 of the swash plate 310. It will be understoodthat the pressure P_(J_piston) is proportional to the angle 342.

Referring now to FIG. 4A, a diagram of a first feedforward controlarchitecture 400 is shown for regulating the hydrostatic differentialpressure pDiff, by controlling a pressure differential of a hydraulicfluid on two sides of a pump control piston 438. The piston position ofa pump control piston 438 is naturally adjusted based on the equilibriumdescribed in previous and following paragraphs. The pump control piston438 and the pump control piston chamber may be non-limiting examples ofthe pump control piston 302 and the pump control piston chamber 306and/or the pump control piston 224 and the pump control piston chamber228 of the hydrostatic unit 200. The control architecture 400 may beimplemented as logic modules stored in memory and executable by aprocessor. Specifically, the different blocks in the controlarchitecture may be conceptually divided into separate modules each withsets of instructions that correspond to control strategies elaboratedupon herein. The pump control piston 438 may control a displacement of ahydrostatic pump of a hydrostatic unit 418, that is fluidly coupled to ahydrostatic motor of the hydrostatic unit 418, as described above inreference to FIG. 2 . The first feedforward control architecture 400 maybe implemented by a controller of a vehicle with an HVT, such as thecontroller 112 of FIG. 1 , to control the hydrostatic unit 418 in afeedforward torque control mode. A processor of the controller (e.g.,the processor 140 of the controller 112 of FIG. 1 ) may executeinstructions stored on a memory of the controller (e.g., the memory 142of the controller 112 of FIG. 1 ) to implement the first feedforwardcontrol architecture 400. In an embodiment, the first feedforwardcontrol architecture 400 uses a non-linear, multi-coefficient model 404to estimate the pressure differential of the hydraulic fluid on the twosides of a pump control piston 438. The model 404 specificallycharacterizes the equilibrium of the pressures acting on the pumpcontrol piston 438.

In one example, the model 404 is derived using the Newton-Euler formula:

J{umlaut over (α)} _(pump) =F(p _(Diff),α_(pump) ,ΔP _(x))  (1)

where J is an inertia of the modelled system, α_(pump) is a variabledisplacement of the hydrostatic pump measured or estimated based on aspeed of the hydrostatic motor, p_(Diff) is a hydrostatic differentialpressure of a hydraulic fluid of the hydrostatic pump, as describedabove in reference to FIGS. 3A-3C, and ΔP_(x) is a pressure differentialbetween a pressure P_(x1) applied to the pump control piston in a firstdirection and a pressure P_(x2) applied to the pump control piston in asecond direction. The differential pressure p_(Diff), pressure P_(x1)and the pressure P_(x2) may be the same as or similar to thedifferential pressure p_(Diff), pressure P_(x1), and the pressure P_(x2)described above in reference to FIGS. 3A-3C. The model 404 implements anequation based on the Newton-Euler formula, where j{umlaut over(α)}_(pump) is eliminated due to a negligible inertia J of movingcomponents of the modelled system (e.g., pump control piston, swashplate, and pump pistons) as follows:

A(n _(engine))p _(Diff) +B(n _(engine))p _(Diff) ² +SW(n_(engine)|_(motor mode))·|α_(pump) |·p _(Diff) +F(n _(engine))·α_(pump)=ΔP _(x)  (2)

In equation (2) A is a first coefficient, B is a second coefficient, SWis a third coefficient, F is a fourth coefficient (also collectivelyreferred to herein as the coefficients), and n_(engine) is a speed ofthe engine (which may be a gain of the pump speed, as the pump speed andthe engine speed are mechanically linked). Prior to operating the HVT,the first coefficient A, the second coefficient B, the third coefficientSW, and the fourth coefficient F may be calibrated. In one example, thecoefficients may be calibrated via an automatic calibration procedureimplemented at a late stage in manufacturing and at predeterminedoperating intervals (e.g., at servicing intervals). During the automaticcalibration procedure, ΔP_(x), p_(Diff) and α_(pump) may be measured orestimated at different engine speeds n_(engine), and the coefficientsmay be calibrated based on a trend of the coefficients. An exampleautomatic calibration procedure is described below in reference to FIG.5 . In one example, the hardware in the pump may be designed to increase(e.g., maximize) the gain between p_(Diff) and ΔP_(x). Further, thecoefficients SW and B and corresponding terms were unexpectedlyidentified using extensive research and development.

The model 404 receives as an input signal a desired hydrostaticdifferential pressure P_(Diff) 402. The desired differential pressureP_(Diff) 402 may be determined by the controller based on an inputsignal from an operator of the vehicle. For example, the operator mayincrease a pressure on an accelerator pedal of the vehicle to increasean amount of torque delivered to one or more wheels of the vehicle, orthe operator may decrease the pressure on an accelerator pedal of thevehicle to decrease the amount of torque delivered to the one or morewheels of the vehicle. In one example, desired differential pressureP_(Diff) 402 may be proportional to a position of the accelerator. Thedifferential pressure P_(Diff) 402 may depend on an angle of a swashplate of the hydrostatic unit, where an increase in a magnitude of theangle of the swash plate may cause the differential pressure P_(Diff)402 to increase, and a decrease in the magnitude of the angle of theswash plate may cause the differential pressure P_(Diff) 402 todecrease. The differential pressure P_(Diff) 402 may be maximized whenthe angle of the swash plate is at a maximum angle, whereby if theoperator presses the accelerator pedal to a fully depressed position,the swash plate is at the maximum angle.

The model 404 also receives as input signals a measured differentialpressure p_(Diff) 424 (e.g., a difference between measurements ofpressure sensors of the hydrostatic unit), and an output of adisplacement calculation block 405, which takes as an input a motorspeed 422 estimated by speed sensors of the motor of the firstfeedforward control architecture 400. The output of the displacementcalculation block 405 is an estimated pump displacement calculated basedon the motor speed 422 and known motor efficiency terms.

An output of the model 404 is a modelled ΔP_(x) 408, which is a desiredΔP_(x) that is a result of the equation (2) above. The model 404 alsooutputs an unmodelled ΔP_(x) to an unmodelled ΔP_(x) block 409. Theunmodelled ΔP_(x) block 409 corrects for model errors (e.g., fromcontrol disturbances due to frictions and uncertainties, etc.),outputting a difference between the unmodelled ΔP_(x) and a differentialpressure of the pump control piston ΔP_(x) 420, where the differentialpressure ΔP_(x) 420 is a measured difference between two pressuresensors arranged at opposite ends of the pump control chamber after aprevious cycle of the feedforward control architecture 400.

In addition to being an input into the model 404, the desiredhydrostatic differential pressure p_(Diff) 402 is also an input into ap_(Diff) error block 406. A second input into the p_(Diff) error block406 is the measured differential pressure p_(Diff) 424. The p_(Diff)error block 406 outputs a difference (e.g., a p_(Diff) error) betweenthe differential pressure p_(Diff) 424 and the desired differentialpressure p_(Diff) 402, which is an input into a p_(Diff) integralcorrection block 412. The p_(Diff) integral correction block 412 mayadjust the p_(Diff) error, correcting for actuation errors, modelerrors, and so forth.

The modelled ΔP_(x) 408 is a first input into a summation block 410. Asecond input into the summation block 410 is a difference between theΔP_(x) 420 and the unmodelled ΔP_(x) of the unmodelled ΔP_(x) block 409,scaled by a gain factor. A third input into the summation block 410 isthe output of the p_(Diff) integral correction block 412. The summationblock 410 sums the first input, the second input, and the third input tooutput a desired ΔP_(x), which is an input of the ΔP_(x) error block414. A second input into the ΔP_(x) error block 414 is the measuredoutput ΔP_(x) 420 of a previous cycle. The output of the ΔP_(x) errorblock 414 is a difference between the measured output ΔP_(x) 420 and thedesired ΔP_(x) 414 (e.g., the summation of the first input, the secondinput, and the third input), which is an input into a valve actuationcorrection block 416. The valve actuation correction block 416translates the desired pressure ΔPx 414 in set-point current for an X1valve 432 and an X2 valve 436, controlled by an X1 valve currentcontroller 430 and an X2 current controller 434, respectively. The valveactuation correction block 416 may also correct for error in valveactuation (e.g., valve hysteresis, flow effect, etc.), where the outputof the valve actuation correction block 416 is outputted to the X1 valvecurrent controller 430 and the X2 current controller 434, respectively,which control the current in solenoids of X1 valve 432 and X2 valve 436.The controlled current is fed into the X1 valve 432 and X2 valve 436,which control the pressures P_(x1) and P_(x2) on the pump control piston438. In other words, The output of block 416 is split and actuated bythe two solenoids of the pressure valves controlling the pressures onthe pump control piston chambers. Thus, in the torque control mode shownin FIG. 4A, the first feedforward control architecture 400 dynamicallyadjusts the ΔP_(x) 420 based on the measured differential pressureP_(Diff) 424 to achieve the target, desired differential pressurep_(Diff) 402 (based on operator input). The adjusted output ΔP_(x) 420may be used by the controller to adjust the pressures P_(x1), and P_(x2)by actuating their respective valves X1 valve 432 and X2 valve 436(e.g., the first pump control piston valve 229 and the second pumpcontrol piston valve 231 of the hydrostatic unit 200 of FIG. 2 ).Concurrently, torque is generated in one or more wheels of the vehicleas a function of the differential pressure p_(Diff) 424, according tothe following equation:

Torque=K·η·p _(Diff)  (3)

where η is a coefficient representing a hydromechanical efficiency ofthe motor 418, and K is a coefficient representing a mechanical gainfrom the motor 418 to the wheels.

Certain control strategies may involve switching between the torque andspeed control modes, embodied via the control architectures 400 and 450illustrated in FIGS. 4A and 4B, respectively. When switching between thetorque and speed control modes, a discontinuity may be perceived by theoperator of the vehicle. In one example, the discontinuity is reduced byincluding junction filters 442, 444, and 446 on inputs of thefeedforward terms. The discontinuity may also be reduced by smartinitialization of an integral term of the p_(Diff) integral correctionblock 412 and/or the displacement integral correction block 456, where astep in the integral term is averted by establishing a functionalrequirement that a contribution of the integral term at a mode switch beequal to a contribution of the integral term at a previous step. To meetthis requirement, the integral term may be reset to a last output valuewhen a mode switch occurs.

Referring now to FIG. 4B, a diagram of a second feedforward controlarchitecture 450 is shown for adjusting a position of the pump controlpiston 438 of the hydrostatic unit 418 to a desired position, bycontrolling a pressure differential of a hydraulic fluid on two sides ofthe pump control piston 438. The pump control piston may be anon-limiting example of the pump control piston 302 and/or the pumpcontrol piston 224 of the hydrostatic unit 200. The pump control piston438 may control a displacement of a hydrostatic pump of the hydrostaticunit that is fluidly coupled to a hydrostatic motor of the hydrostaticunit, as described above in reference to FIG. 2 . The second feedforwardcontrol architecture 450 may be implemented by a controller of a vehiclewith an HVT, such as the controller 112 of FIG. 1 , to control thehydrostatic unit 418 in a feedforward speed control mode (e.g., ratherthan the torque control mode of FIG. 4A). A processor of the controller(e.g., the processor 140 of the controller 112 of FIG. 1 ) may executeinstructions stored on a memory of the controller (e.g., the memory 142of the controller 112 of FIG. 1 ) to implement the second feedforwardcontrol architecture 450.

The second feedforward control architecture 450 may be substantiallysimilar to the first feedforward control architecture 400 describedabove in reference to FIG. 4A, and may share many of the components ofthe first feedforward control architecture 400, which are similarlynumbered for clarity. In particular, the second feedforward controlarchitecture 450 uses the same model 404 to estimate pressures andforces of the hydrostatic unit, with the same coefficients andparameters. As in the torque control mode of the first feedforwardcontrol architecture 400, in the speed control mode, the model 404 maybe used to adjust the ΔP_(x) 420 based on a desired pump displacement452. The desired pump displacement 452 is an input into the model 404,along with the differential pressure p_(Diff) 424 and the output of thedisplacement calculation block 405. The p_(Diff) error block 406 and thep_(Diff) integral correction block 412 of the first feedforward controlarchitecture 400 of FIG. 4A are replaced by a displacement error block454 and a displacement integral correction block 456, respectively. Thedesired pump displacement 452 is an input into the displacement errorblock 454, in parallel with the model 404. An additional input into thedisplacement error block 454 is the output of the displacementcalculation block 405. The displacement error block 454 outputs adifference (e.g., a displacement error) between the desired pumpdisplacement 452 and the output of the displacement calculation block405, which is an input into the displacement integral correction block456. The displacement integral correction block 456 adjusts thedisplacement error, correcting for actuation errors, model errors, andso forth. The output of the displacement integral correction block 456is a third input into the summation block 410, along with the modelledΔP_(x) 408 (the first input into a summation block 410) and thedifference between the ΔP_(x) 420 and the unmodelled ΔP_(x) scaled bythe gain factor (the second input into the summation block 410). Asdescribed above in reference to FIG. 4A, the summation block 410 sumsthe first input, the second input, and the third input to output adesired ΔP_(x), which is inputted into ΔP_(x) error block 414. Theremaining components of the second feedforward control architecture 450are identical to the first feedforward control architecture 450 of FIG.4A.

Thus, in the speed control mode shown in FIG. 4B, the second feedforwardcontrol architecture 450 dynamically adjusts the pump displacement byminimizing a difference between the output of the displacementcalculation block 405 based on the motor speed 422 of the motor and thetarget, desired pump displacement 452, thereby generating an adjustedΔP_(x) 420 that may be used by the controller to adjust the pressuresP_(x1), and P_(x2) by actuating their respective valves (e.g., X1 valve432 and X2 valve 436). In one example, the desired pump displacement 452is generated by the controller based on operator input via an inputdevice (e.g., one of the input devices 114 of the HVT control system 100of FIG. 1 ), for example, when engaging a first forward drive clutch orreverse drive clutch via an FNR lever. Concurrently, the motor speed iscontrolled as a function of the pump displacement α_(pump), according tothe following equation:

Speed=K·η·Displ  (3)

Where η is the coefficient representing the volumetric efficiency of thehydrostatic unit, and K is the coefficient representing a mechanicalgain from the motor 418 to the wheels.

Referring now to FIG. 5 , a flowchart of an exemplary method 500 isshown for calibrating a first coefficient A, a second coefficient B, athird coefficient SW, and a fourth coefficient F (herein, thecoefficients) of a non-linear, multi-coefficient model (herein, themodel), where the model is implemented within a feed-forward controlarchitecture for controlling a pump control piston of a hydrostatic unitof an HVT, such as the pump control piston 224 of the hydrostatic unit200 of FIG. 2 . The first coefficient A, the second coefficient B, thethird coefficient SW, and the fourth coefficient F may be the same as orsimilar to the first coefficient A, the second coefficient B, the thirdcoefficient SW, and the fourth coefficient F of the non-linear,multi-coefficient model 404 of FIGS. 4A and 4B. In an embodiment,operations of method 500 may be stored in non-transitory memory andexecuted by a processor, such as memory 142 and processor 140 of the HVTcontrol system 100 of FIG. 1 , respectively, during operation of an HVTsuch as the HVT 101 of the HVT control system 100 of FIG. 1 . Asdescribed above in reference to FIG. 4A, during the automaticcalibration procedure, ΔP_(x), p_(Diff), α_(pump) may be measured orestimated n times at different engine speeds n_(engine), and thecoefficients may be calibrated based on a trend of the coefficients, inaccordance with the modified Newton-Euler formula described above inreference to FIG. 4A:

A(n _(engine))p _(Diff) +B(n _(engine))p _(Diff) ² +SW(n_(engine)|_(motor mode))·|α_(pump) |·p _(Diff) +F(n _(engine))·α_(pump)=ΔP _(x)  (4)

At 502, method 500 includes estimating and/or measuring vehicleoperating conditions. Vehicle operating conditions may be estimatedbased on one or more outputs of various sensors of the vehicle (e.g.,such as an oil temperature sensor, engine velocity or wheel velocitysensor, torque sensor, swash plate angle sensor, pressure sensor, etc.,as described above in reference to the HVT control system 100 of FIG. 1and the hydrostatic unit 200 of FIG. 2 ). Vehicle operating conditionsmay include engine velocity and load, vehicle velocity, transmission oiltemperature, exhaust gas flow rate, mass air flow rate, coolanttemperature, coolant flow rate, engine oil pressures (e.g., oil gallerypressures), operating modes of one or more intake valves and/or exhaustvalves, electric motor velocity, battery charge, engine torque output,vehicle wheel torque, etc. Estimating and/or measuring vehicle operatingconditions may include determining whether the vehicle is being poweredby an engine or an electric motor (e.g., the power source 102 of the HVTcontrol system of FIG. 1 ).

At 504, method 500 includes disengaging a plurality of drive clutches ofthe HVT (e.g., the reverse drive clutch 160, first forward drive clutch162, and second forward drive clutch 164 of the HVT control system 100of FIG. 1 ). By disengaging the plurality of drive clutches of the HVT,a differential pressure p_(Diff) may be reduced to null due to thehydraulic motor of the HVT generating a torque that is close to null,where a negligible amount of torque is generated to counteract aninternal friction of the hydraulic motor and/or attached gears,bearings, and clutches. As p_(Diff) is (approximately) linearlyproportional to motor torque, p_(Diff) reduces to near 0. When thedifferential pressure p_(Diff) is reduced to null, the first three termsof the modified Newton-Euler formula may be eliminated, leaving thefollowing equation:

F(n _(engine))·α_(pump) =ΔP _(x)  (5)

which may be reformulated as the following “no load curve” equation:

$\begin{matrix}{F_{(n_{engine})} = \frac{\Delta P_{x}}{\alpha_{pump}}} & (6)\end{matrix}$

At 506, method 500 includes calibrating the fourth coefficient F of themodel at a plurality of different engine speeds (e.g., three or moreengine speeds). To determine the coefficient F, the clutches in thetransmission are disengaged which results in a no load condition wherep_(Diff) is zero. In one example, the calibration curve for thecoefficient F, referred to as the “no load curve,” is solved for byswiveling the pump under the no load condition Referring briefly to FIG.6 , a use-case three dimensional (3D) surface plot 600 is shown of therelationship between ΔP_(x), the pump displacement a_(pump) and thedifferential pressure p_(Diff) at a selected engine speed as dictated byequation (4). The relationships shown in FIG. 6 correspond to anexemplary variable displacement pump and the numerical values for thevariables are therefore for illustration purposes only and are not meantto be limiting. A no load curve 602 where the differential pressurep_(Diff) is 0 and the pump is swivelled through its angular range isindicated on the surface plot. As previously mentioned, the calibrationcurve corresponding to the coefficient F may be repeated at a pluralityof engine speeds.

At 508, method 500 includes operating the HVT in a blocking condition.In one example, the blocking condition is generated by engaging twoclutches of the HVT, thereby locking a transmission output of the HVT(e.g., whereby no torque is delivered to one or more wheels of thevehicle). For example, if there are three drive clutches, a reversedrive clutch and a first forward drive clutch (e.g., the first forwarddrive clutch 162 and the reverse drive clutch 160 of FIG. 1 ) may beengaged. During operation in the blocking condition, an increase ordecrease in ΔP_(x) generates a corresponding increase or decrease in thedifferential pressure p_(Diff).

At 510, method 500 includes operating a hydrostatic unit of the HVT in apump mode during the blocking condition. For example, the hydrostaticunit may include a hydrostatic pump and a hydrostatic motor (e.g., thehydrostatic pump 202 and the hydrostatic motor 204 of the hydrostaticunit 200 of FIG. 2 ), where the hydrostatic unit may be operated in apump mode or a motor mode by adjusting an angle of a swash plate of thehydrostatic unit, as described above in reference to FIG. 2 . Thehydrostatic unit may be operated in the pump mode by adjusting the swashplate to have a positive angle with respect to a pump control piston(e.g., the pump control piston 224 of FIG. 2 ) whereby a rotation of apump carrier of the hydrostatic pump (e.g., the pump carrier 208 of FIG.2 ) causes the hydrostatic motor to rotate in a first direction, or thehydrostatic unit may be operated in the motor mode by adjusting theswash plate to have a negative angle with respect to a pump controlpiston, whereby the rotation of a pump carrier causes the hydrostaticmotor to rotate in a second direction. Since the terms of the equation(4) with coefficients A and B apply to the pump mode, and the term ofthe equation (4) with the coefficient SW applies to the motor mode, A,B, and SW may be calculated by selectively operating the hydrostaticunit of the HVT in the pump mode or the motor mode to alternatelyeliminate and solve for the relevant term(s) of the equation (4). Duringoperation in the pump mode, ΔP_(x) is increased to increase P_(Diff),while during operation in the motor mode, ΔP_(x) is decreased todecrease P_(Diff).

Since the third coefficient SW of the model only applies in a “motormode”, by operating the HVT in the pump mode, an effect of thecoefficient SW may be nullified, allowing the third term of the modifiedNewton-Euler formula above to be eliminated, leaving the following “loadcurve” equation:

A(n _(engine))p _(Diff) +B(n _(engine))p _(Diff) ² +F(n_(engine))·α_(pump) =ΔP _(x)  (7)

which may be reformulated as follows:

A(n _(engine))p _(Diff) +B(n _(engine))p _(Diff) ² =ΔP _(x) −F(n_(engine))·α_(pump)  (8)

where F is known.

At 512, method 500 includes calibrating the first coefficient A and thesecond coefficient B of the modified Newton-Euler formula after thecoefficient F is known, where the coefficients A and B are the linearand quadratic coefficients, respectively, of the “load curve” parabola.As described above, method 500 includes increasing the ΔP_(x) (e.g., byoperating in the pump mode) and calibrating the coefficients A and B ata number n of different engine speeds, where a trend in the coefficientsA and B are used to calibrate the coefficients A and B. Referring againto FIG. 6 , the load curve representing the calibration curve ofcoefficients A and B is indicated at 606. As shown, the load curve 606is determined by increasing ΔP_(x) and p_(Diff) relative to the pumpdisplacement α_(pump). This calibration curve procedure may again berepeated at different engine speeds.

At 514, method 500 includes switching from the pump mode to a motormode, and operating the HVT in the motor mode to decrease ΔP_(x) and thedifferential pressure p_(Diff) once the coefficients A, B, and F havebeen determined, where the modified Newton-Euler formula is now invertedin a “drag curve” equation as follows:

$\begin{matrix}{{SW}_{({n_{engine}❘{{motor}{mode}}})} = {{\Delta P_{x}} - {A_{(n_{engine})}p_{Diff}} + {B_{(n_{engine})}p_{Diff}^{2}} - {F_{(n_{engine})} \cdot {\alpha_{pump}/{❘\alpha_{pump}❘}} \cdot p_{Diff}}}} & (9)\end{matrix}$

At 516, method 500 includes solving for and calibrating the thirdcoefficient SW of the model. As described above, method 500 includesdecreasing the ΔP_(x) and calibrating the coefficient SW at a number nof different engine speeds, where a trend in the coefficient SW is usedto calibrate the coefficient SW. Referring again to FIG. 6 , thecoefficient curve SW (referred to as the drag curve) along theexperimental map is indicated at 604. To elaborate, the drag curve maybe found by inverting equation (9) and decreasing ΔP_(x) and p_(Diff)relative to the pump displacement α_(pump).

Thus, an automatic procedure for calibrating the four coefficients isproposed, where each of the four coefficients are calibrated in turn.Clutches are first disengaged to solve for a coefficient F; clutches aresubsequently engaged to increase ΔP_(x) while operating the hydrostaticunit in the pump mode to solve for the coefficients A and B,respectively; and then the hydrostatic unit is operated in the motormode to decrease ΔP_(x) to determine the coefficient SW. Values of thecoefficients are determined in n trials with n different engine speeds,and the coefficients are calibrated based on trends in the values of thecoefficients over the n trials. After the calibration procedure, acommon non-linear, multi-coefficient model (e.g., the modifiedNewton-Euler formula) may be used as a block in a first controlarchitecture for controlling the hydrostatic unit in a torque controlmode, and/or as a block in a second control architecture for controllingthe hydrostatic unit in a transmission speed control mode. In this way,the differential pressure p_(Diff) may be controlled by compensating thepump displacement in the torque control mode, or the pump displacementmay be controlled by compensating the differential pressure p_(Diff) inthe speed control mode. This allows the HVT to efficiently switchbetween the speed and torque control modes, and minimize a discontinuitybetween the speed and torque control modes when switching. As a result,a performance of the vehicle during operation may be increased and awear on parts of the transmission may be decreased. An additionaladvantage of the method is that the torque control mode and the speedcontrol mode share a common model and a common procedure forcalibration, which may reduce a use of computational resources and/orreduce a latency when switching between control modes.

The technical effect of using a common, non-linear multi-coefficientmodel and calibration procedure for controlling a hydrostatic unit wherethe differential pressure p_(Diff) is controlled in a torque controlmode and a pump displacement α_(pump) is controlled in a speed controlmode is that an accuracy of a feedforward control model may beincreased, an efficiency of switching between the speed and torquecontrol modes may be maximized, and a discontinuity between the speedand torque control modes when switching may be reduced (e.g., avoided).

The disclosure also provides support for a control method for ahydrostatic unit of a hydromechanical variable transmission (HVT),comprising: controlling the hydrostatic unit via a feedforward controlarchitecture including a non-linear, multi-coefficient model, whereinthe hydrostatic unit comprises a hydrostatic pump and a hydrostaticmotor. In a first example of the method, a displacement of thehydrostatic pump is controlled by an angle of a swash plate of thehydrostatic unit, the angle of the swash plate is controlled by a pumpcontrol piston of the hydrostatic unit, and an output of the feedforwardcontrol architecture is a change in pressure differential of a hydraulicfluid on two sides of a pump control piston. In a second example of themethod, optionally including the first example, controlling thehydrostatic unit further comprises switching between operating the HVTin a torque control mode and operating the HVT in a speed control modeof the hydrostatic unit based on vehicle operating conditions, and thesame non-linear, multi-coefficient model is used for both the torquecontrol mode and the speed control mode. In a third example of themethod, optionally including one or both of the first and secondexamples, the HVT is engaged in the torque control mode and the HVT isdisengaged in the speed control mode. In a fourth example of the method,optionally including one or more or each of the first through thirdexamples, the hydrostatic unit is operated in the speed control mode ata first engagement of a gear of the HVT, to synchronize a desired clutchof the HVT. In a fifth example of the method, optionally including oneor more or each of the first through fourth examples, the hydrostaticunit is operated in the speed control mode during a freewheel conditionto follow a free output speed of the HVT in preparation of a fastre-engagement of a clutch of the HVT. In a sixth example of the method,optionally including one or more or each of the first through fifthexamples, operating the HVT in the torque control mode includescontrolling a differential pressure of the hydrostatic pump bycompensating a pump displacement of the hydrostatic pump, and operatingthe HVT in the speed control mode includes controlling the pumpdisplacement by compensating the differential pressure. In a seventhexample of the method, optionally including one or more or each of thefirst through sixth examples, during operation in the torque controlmode, an input into the feedforward control architecture is a desireddifferential pressure, and during operation in the speed control mode,an input into the feedforward control architecture is a desired pumpdisplacement. In an eighth example of the method, optionally includingone or more or each of the first through seventh examples, thenon-linear, multi-coefficient model has a plurality of coefficients thatare calibrated at predetermined time intervals via an automaticcalibration procedure that operates one or more clutches and thehydrostatic unit to determine the plurality of coefficients, the methodfurther comprising controlling the pump control piston via thefeedforward control architecture using the non-linear, multi-coefficientmodel with the calibrated coefficients. In a ninth example of themethod, optionally including one or more or each of the first througheighth examples, one or more inputs to the feedforward controlarchitecture include junction filters to reduce a discontinuity whenswitching between the torque control mode and the speed control mode. Ina tenth example of the method, optionally including one or more or eachof the first through ninth examples, the automatic calibration procedureincludes: solving for the four coefficients at a plurality of enginespeeds, identifying a trend in each coefficient of the four coefficientsover the plurality of engine speeds, and calibrating each coefficient ofthe four coefficients based on the trend of each coefficient.

The disclosure also provides support for a control system for ahydromechanical variable transmission (HVT) of a vehicle, comprising: ahydrostatic unit including a hydrostatic pump and a hydrostatic motorrotationally coupled in parallel with a planetary gear set, a controllerincluding a processor and instructions stored on a non-transitory memoryof the controller that when executed cause the controller to: inresponse to a request to control a torque of the hydrostatic motor,implement a first feedforward control architecture including anon-linear, multi-coefficient model, where inputs into the non-linear,multi-coefficient model include a desired differential pressure of thehydrostatic pump, a measured differential pressure of the hydrostaticpump, and an estimated displacement of the hydrostatic pump, in responseto a request to control a speed of the hydrostatic motor, implement asecond feedforward control architecture including the non-linear,multi-coefficient model, where inputs into the non-linear,multi-coefficient model include a desired displacement of thehydrostatic pump, a measured differential pressure of the hydrostaticpump, and an estimated displacement of the hydrostatic pump, and adjusta position of a pump control piston of the hydrostatic unit based on anoutput of the non-linear, multi-coefficient model, where the output is adesired differential pressure of a hydraulic fluid on two sides of thepump control piston. In a first example of the system, the request tocontrol the torque of the hydrostatic motor is in response to anengagement of a gear of the HVT, and the request to control the speed ofthe hydrostatic motor is in response to a disengagement of a gear of theHVT. In a second example of the system, optionally including the firstexample, the first feedforward control architecture and the secondfeedforward control architecture include junction filters for the inputsof the non-linear, multi-coefficient model. In a third example of thesystem, optionally including one or both of the first and secondexamples, an integral term of the first feedforward control architectureor the second feedforward control architecture is reset to a last outputvalue when switching between a speed control mode and a torque controlmode. In a fourth example of the system, optionally including one ormore or each of the first through third examples, the non-linear,multi-coefficient model implements a polynomial non-linear equation withfour coefficients based on the Newton-Euler formula. In a fifth exampleof the system, optionally including one or more or each of the firstthrough fourth examples, the four coefficients are calibrated by anautomatic calibration procedure including: reducing a differentialpressure of the hydrostatic unit to zero to solve for a fourthcoefficient of the four coefficients, knowing the fourth coefficient,increasing the differential pressure to solve for a first coefficientand a second coefficient of the four coefficients, knowing the fourthcoefficient, the first coefficient and the second coefficient,decreasing the differential pressure to solve for a third coefficient ofthe four coefficients. In a sixth example of the system, optionallyincluding one or more or each of the first through fifth examplescomprising further instructions stored in memory that when executed,cause the controller to: disengage clutches of the planetary gear set toreduce the reduce the differential pressure to zero, engage clutches ofthe planetary gear set and operate the hydrostatic unit in a pump modeto increase the differential pressure, and engage clutches of theplanetary gear set and operate the hydrostatic unit in a motor mode todecrease the differential pressure. In a seventh example of the system,optionally including one or more or each of the first through sixthexamples comprising further instructions stored in memory that whenexecuted, cause the controller to: solve for the four coefficients at aplurality of engine speeds, and based on a trend in each coefficient ofthe four coefficients over the plurality of engine speeds, calibrateeach coefficient based on the trend of each coefficient.

The disclosure also provides support for a control method for ahydrostatic unit of a hydromechanical variable transmission (HVT) of avehicle, comprising: in a first mode of operation, using a firstfeedforward control architecture including a non-linear,multi-coefficient model to adjust a pressure of a hydraulic fluid on apump control piston of the hydrostatic unit following a torquereference, and in a second mode of operation, using a second feedforwardcontrol architecture including the non-linear, multi-coefficient modelto adjust a pressure of a hydraulic fluid on a pump control piston ofthe hydrostatic unit following a speed reference, wherein thehydrostatic unit switches between the first mode of operation and thesecond mode of operation dynamically based on input from an operator ofthe vehicle.

While various embodiments have been described above, it should beunderstood that they have been presented by way of example, and notlimitation. It will be apparent to persons skilled in the relevant artsthat the disclosed subject matter may be embodied in other specificforms without departing from the spirit of the subject matter. Theembodiments described above are therefore to be considered in allrespects as illustrative, not restrictive.

Note that the example control and estimation routines included hereincan be used with various powertrain and/or vehicle systemconfigurations. The control methods and routines disclosed herein may bestored as executable instructions in non-transitory memory and may becarried out by the control system including the controller incombination with the various sensors, actuators, and other transmissionand/or vehicle hardware. Further, portions of the methods may bephysical actions taken in the real world to change a state of a device.The specific routines described herein may represent one or more of anynumber of processing strategies such as event-driven, interrupt-driven,multi-tasking, multi-threading, and the like. As such, various actions,operations, and/or functions illustrated may be performed in thesequence illustrated, in parallel, or in some cases omitted. Likewise,the order of processing is not necessarily required to achieve thefeatures and advantages of the example examples described herein, but isprovided for ease of illustration and description. One or more of theillustrated actions, operations and/or functions may be repeatedlyperformed depending on the particular strategy being used. Further, thedescribed actions, operations and/or functions may graphically representcode to be programmed into non-transitory memory of the computerreadable storage medium in the vehicle and/or transmission controlsystem, where the described actions are carried out by executing theinstructions in a system including the various hardware components incombination with the electronic controller. One or more of the methodsteps described herein may be omitted if desired.

It will be appreciated that the configurations and routines disclosedherein are exemplary in nature, and that these specific examples are notto be considered in a limiting sense, because numerous variations arepossible. For example, the above technology can be applied topowertrains that include different types of propulsion sources includingdifferent types of electric machines, internal combustion engines,and/or transmissions. The subject matter of the present disclosureincludes all novel and non-obvious combinations and sub-combinations ofthe various systems and configurations, and other features, functions,and/or properties disclosed herein.

The following claims particularly point out certain combinations andsub-combinations regarded as novel and non-obvious. These claims mayrefer to “an” element or “a first” element or the equivalent thereof.Such claims should be understood to include incorporation of one or moresuch elements, neither requiring nor excluding two or more suchelements. Other combinations and sub-combinations of the disclosedfeatures, functions, elements, and/or properties may be claimed throughamendment of the present claims or through presentation of new claims inthis or a related application. Such claims, whether broader, narrower,equal, or different in scope to the original claims, also are regardedas included within the subject matter of the present disclosure.

As used herein, the term “substantially” is construed to mean plus orminus five percent of the range, unless otherwise specified.

1. A control method for a hydrostatic unit of a hydromechanical variabletransmission (HVT), comprising: controlling the hydrostatic unit via afeedforward control architecture including a non-linear,multi-coefficient model; wherein the hydrostatic unit comprises ahydrostatic pump and a hydrostatic motor.
 2. The method of claim 1,wherein a displacement of the hydrostatic pump is controlled by an angleof a swash plate of the hydrostatic unit, the angle of the swash plateis controlled by a pump control piston of the hydrostatic unit, and anoutput of the feedforward control architecture is a change in pressuredifferential of a hydraulic fluid on two sides of a pump control piston.3. The method of claim 2, wherein controlling the hydrostatic unitfurther comprises switching between operating the HVT in a torquecontrol mode and operating the HVT in a speed control mode of thehydrostatic unit based on vehicle operating conditions, and the samenon-linear, multi-coefficient model is used for both the torque controlmode and the speed control mode.
 4. The method of claim 3, wherein theHVT is engaged in the torque control mode and the HVT is disengaged inthe speed control mode.
 5. The method of claim 4, wherein thehydrostatic unit is operated in the speed control mode at a firstengagement of a gear of the HVT, to synchronize a desired clutch of theHVT.
 6. The method of claim 4, wherein the hydrostatic unit is operatedin the speed control mode during a freewheel condition to follow a freeoutput speed of the HVT in preparation of a fast re-engagement of aclutch of the HVT.
 7. The method of claim 3, wherein operating the HVTin the torque control mode includes controlling a differential pressureof the hydrostatic pump by compensating a pump displacement of thehydrostatic pump, and operating the HVT in the speed control modeincludes controlling the pump displacement by compensating thedifferential pressure.
 8. The method of claim 7, wherein duringoperation in the torque control mode, an input into the feedforwardcontrol architecture is a desired differential pressure, and duringoperation in the speed control mode, an input into the feedforwardcontrol architecture is a desired pump displacement.
 9. The method ofclaim 8, wherein the non-linear, multi-coefficient model has a pluralityof coefficients that are calibrated at predetermined time intervals viaan automatic calibration procedure that operates one or more clutchesand the hydrostatic unit to determine the plurality of coefficients, themethod further comprising controlling the pump control piston via thefeedforward control architecture using the non-linear, multi-coefficientmodel with the calibrated coefficients.
 10. The method of claim 3,wherein one or more inputs to the feedforward control architectureinclude junction filters to reduce a discontinuity when switchingbetween the torque control mode and the speed control mode.
 11. Themethod of claim 9, wherein the automatic calibration procedure includes:solving for the four coefficients at a plurality of engine speeds;identifying a trend in each coefficient of the four coefficients overthe plurality of engine speeds; and calibrating each coefficient of thefour coefficients based on the trend of each coefficient.
 12. A controlsystem for a hydromechanical variable transmission (HVT) of a vehicle,comprising: a hydrostatic unit including a hydrostatic pump and ahydrostatic motor rotationally coupled in parallel with a planetary gearset; a controller including a processor and instructions stored on anon-transitory memory of the controller that when executed cause thecontroller to: in response to a request to control a torque of thehydrostatic motor, implement a first feedforward control architectureincluding a non-linear, multi-coefficient model, where inputs into thenon-linear, multi-coefficient model include a desired differentialpressure of the hydrostatic pump, a measured differential pressure ofthe hydrostatic pump, and an estimated displacement of the hydrostaticpump; in response to a request to control a speed of the hydrostaticmotor, implement a second feedforward control architecture including thenon-linear, multi-coefficient model, where inputs into the non-linear,multi-coefficient model include a desired displacement of thehydrostatic pump, a measured differential pressure of the hydrostaticpump, and an estimated displacement of the hydrostatic pump; and adjusta position of a pump control piston of the hydrostatic unit based on anoutput of the non-linear, multi-coefficient model, where the output is adesired differential pressure of a hydraulic fluid on two sides of thepump control piston.
 13. The control system of claim 12, wherein therequest to control the torque of the hydrostatic motor is in response toan engagement of a gear of the HVT, and the request to control the speedof the hydrostatic motor is in response to a disengagement of a gear ofthe HVT.
 14. The control system of claim 12, wherein the firstfeedforward control architecture and the second feedforward controlarchitecture include junction filters for the inputs of the non-linear,multi-coefficient model.
 15. The control system of claim 12, wherein anintegral term of the first feedforward control architecture or thesecond feedforward control architecture is reset to a last output valuewhen switching between a speed control mode and a torque control mode.16. The control system of claim 12, wherein the non-linear,multi-coefficient model implements a polynomial non-linear equation withfour coefficients based on the Newton-Euler formula.
 17. The controlsystem of claim 16, wherein the four coefficients are calibrated by anautomatic calibration procedure including: reducing a differentialpressure of the hydrostatic unit to zero to solve for a fourthcoefficient of the four coefficients; knowing the fourth coefficient,increasing the differential pressure to solve for a first coefficientand a second coefficient of the four coefficients; knowing the fourthcoefficient, the first coefficient and the second coefficient,decreasing the differential pressure to solve for a third coefficient ofthe four coefficients.
 18. The control system of claim 17, comprisingfurther instructions stored in memory that when executed, cause thecontroller to: disengage clutches of the planetary gear set to reducethe reduce the differential pressure to zero; engage clutches of theplanetary gear set and operate the hydrostatic unit in a pump mode toincrease the differential pressure; and engage clutches of the planetarygear set and operate the hydrostatic unit in a motor mode to decreasethe differential pressure.
 19. The control system of claim 16,comprising further instructions stored in memory that when executed,cause the controller to: solve for the four coefficients at a pluralityof engine speeds; and based on a trend in each coefficient of the fourcoefficients over the plurality of engine speeds, calibrate eachcoefficient based on the trend of each coefficient.
 20. A control methodfor a hydrostatic unit of a hydromechanical variable transmission (HVT)of a vehicle, comprising: in a first mode of operation, using a firstfeedforward control architecture including a non-linear,multi-coefficient model to adjust a pressure of a hydraulic fluid on apump control piston of the hydrostatic unit following a torquereference; and in a second mode of operation, using a second feedforwardcontrol architecture including the non-linear, multi-coefficient modelto adjust a pressure of a hydraulic fluid on a pump control piston ofthe hydrostatic unit following a speed reference; wherein thehydrostatic unit switches between the first mode of operation and thesecond mode of operation dynamically based on input from an operator ofthe vehicle.